This invention relates to a turbo compressor having a rotor supported by a magnetic bearing and a method of controlling the turbo compressor, and more particularly, to a high-pressure type of single-shaft multistage centrifugal compressor.
As shown in FIG. 3, a single-shaft multi-stage centrifugal compressor includes a rotor 2 having a plurality of impellers 10 are attached thereto, with the rotor 20 and impellers 10 being incorporated in a casing 1 and, with the rotor 10 being supported by oil bearings 7. The rotor 2 is rotated by a motor to draw a gas being compressed through a nozzle 4. Driving energy is thereby applied to the gas by the impeller 10, and the gas is thereby increased in pressure and is discharged through a nozzle 5.
The pressure of the gas is higher at a compression stage closer to the rear portion of the compressor, and the interstage labyrinth diameter is smaller than the mouth labyrinth diameter. Therefore the static pressure acts on one impeller as an axial force in the direction from the discharge side to the suction side. To reduce this thrust force to a magnitude within the range of an allowable load capacity of the thrust bearing corresponding to the axial-end diameter of the rotor, a balancing drum 11 is provided at an axial end of the final-stage impeller, and a chamber 12 communicating with the suction side is formed outside the balancing drum 11.
In this type of turbo compressors using oil bearings, the friction loss of the bearings is generally several tens of kilowatts. If the above-mentioned oil bearings are replaced with non-contact type magnetic bearings, the running cost of the compressor can be reduced and the lubrication system can be maintenance-free.
As a turbo compressor using a balancing drum, a compressor disclosed in Japanese Patent Unexamined Publication No. 62-258195 is known.
The load capacity of magnetic bearings, however, is on the order of about 1/10 of that of oil bearings (ordinarily, several kg/cm.sup.2 or less in terms of surface pressure), and a problem of an increase in the size of a thrust bearing portion is therefore encountered. That is, if a magnetic bearing is used, the diameter of a journal bearing portion is increased in comparison with oil bearings and the inside diameter of the thrust bearing is also increased, because the load capacity of the magnetic bearing is small. Also there is a need to increase the area of a thrust collar, i.e., the diameter of the thrust collar in comparison with oil bearings. If the thrust bearing is disposed outside of the journal bearing and if the thrust collar is large, a problem of a natural oscillation frequency of the compressor rotor in a high-order mode is reduced close to the number of revolutions in steady operation arises. If the thrust bearing is disposed inside the journal bearing, a problem of a further increase in the diameter of the shaft sealing device, a problem of the allowable peripheral speed of the shaft sealing device being exceeded, and a problem of difficulty in constructing the compressor are encountered.
A thrust force acting on the compressor shaft is calculated from a static pressure on an impeller portion, fluid force caused by a gas flow, leaks through gaps, and so on. However, the thrust force is changed due to the difference between design conditions and operating conditions or by a change in these conditions or a change with time (such as wear of labyrinth seals). Therefore the reliability of the bearing system cannot be satisfactorily high unless a thrust bearing having a sufficiently large load capacity is selected. For this reason, a large thrust bearing having such a large capacity is required in the conventional compressors.